1. Field of the Invention
The present invention generally relates to a control system for an internal combustion engine of cylinder injection type (also known as the direct fuel injection type engine) in which fuel is directly injected into engine cylinders to undergo combustion therein through spark ignition. More specifically, the invention is concerned with a control system for the cylinder injection type internal combustion engine which is capable of ensuring a significant reduction of harmful components contained in the exhaust gas of the engine with high efficiency while realizing improved combustion performance or drivability of the engine.
2. Description of Related Art
In general, with the spark ignition type internal combustion engine in which fuel is injected into an intake manifold for charging a uniform gas mixture into engine cylinders, a relatively high output power or torque can be generated. However, the internal combustion engine (hereinafter also referred to simply as the engine) of this type suffers a problem that the output torque thereof changes rather remarkably in dependence on the air-fuel ratio (hereinafter also referred to simply as the A/F ratio), providing thus difficulty in carrying out the control of output torque generated by the engine.
In the conventional intake manifold injection type engine (indirect fuel injection type engine) known heretofore, the air-fuel ratio can remain substantially constant, whereby the relation between an opening degree of a throttle valve and the output torque of the engine can be determined substantially definitely. For this reason, the combustion parameters such as the ignition timing and others have been determined or set definitely in dependence on the intake air flow rate (hereinafter also referred to as the intake air quantity).
By contrast, in the case of the cylinder injection type internal combustion engine, the output torque of the engine changes in dependence on the air-fuel ratio even for a same opening degree of the throttle valve. Accordingly, there arises the necessity of setting the combustion parameters such as the ignition timing and the fuel injection timing among others to optimal values in dependence on the engine load, the air-fuel ratio and others by controlling appropriately the throttle valve opening degree and the air-fuel ratio.
For having better understanding of the principle underlying the present invention, technical background thereof will be described below in some detail. FIG. 23 is a schematic diagram showing generally an arrangement of a conventional control system for a cylinder injection type internal combustion engine known heretofore. Referring to FIG. 23, an engine 1 constituting a major part of the internal combustion engine system is provided with an intake pipe 1a for introducing the intake air into the engine 1 and an exhaust pipe 1b for discharging the exhaust gas resulting from the combustion of the air-fuel mixture.
An air flow sensor 2 for detecting a flow rate or quantity Qa of the intake air flow fed to the engine 1 as indicated by an arrow is installed at an upstream location of the intake pipe la. Further installed within the intake pipe 1a is a throttle valve 3 for regulating or adjusting the intake air flow rate or quantity Qa, and a throttle position sensor 4 for detecting an opening degree .theta. of the throttle valve 3 is provided in association with the throttle valve 3.
Installed at a downstream location of the intake pipe 1a,i.e., at a location immediately preceding to the engine 1 is a surge tank 5. On the other hand, an air-fuel ratio sensor 6 which may be constituted by a linear type O.sub.2 -sensor is provided in the exhaust pipe 1b to serve for detecting an actual air-fuel (A/F) ratio F prevailing within the exhaust pipe 1b. Parenthetically, this air-fuel ratio may change within a range of e.g. 10 to 50.
A throttle valve actuator which serves as the intake air quantity regulating means 7 for adjusting or regulating the throttle valve opening degree .theta. is provided in association with the throttle valve 3. The throttle valve actuator 7 may be comprised of, for example, a stepping motor for operating rotatively the throttle valve 3 to thereby regulate the rate or quantity Qa of the intake air flowing through the intake pipe 1a.
Installed within each of the cylinders of the engine 1 is a spark plug 8 at which electric spark discharge takes place for igniting the air-fuel mixture within the combustion chamber of the cylinder. To this end, a distributor 9 is provided for supplying a high voltage distributively to the individual spark plugs 8 in synchronism with an ignition timing.
An ignition coil 10 realized in the form of a transformer having primary and secondary windings. The high voltage for ignition is generated across the secondary winding of the ignition coil 10 upon every interruption of the primary current flowing through the primary winding. The high voltage is then supplied to the distributor 9. Provided in association with the ignition coil 10 is an ignitor 11 which is constituted by a power transistor for interrupting the current flowing through the primary winding of the ignition coil 10 in conformance with the ignition timing for the individual engine cylinders.
The spark plug 8, the distributor 9, the ignition coil 10 and the ignitor 11 cooperate to constitute an ignition system for igniting the air-fuel mixture within the individual cylinders of the engine 1.
An ECU (Electronic Control Unit) 12 which is in charge of controlling the engine system as a whole includes a microcomputer for determining arithmetically control parameters for various actuators employed for the combustion control of the engine 1 on the basis of information detected by various sensors (i.e., information concerning the operation state of the engine 1) to thereby issue driving signals indicative of the control parameters to the relevant actuators.
As the control parameter signals, there may be mentioned an intake-air flow control signal A for the throttle valve actuator 7, an ignition signal G for the ignitor 11 (ignition system), a fuel injection signal J for the fuel injection valve (i.e., injector) 13, an EGR (Exhaust Gas Recirculation) control signal E for an EGR regulating valve 17 and a swirl control signal B for the swirl valve actuator (swirl rate regulating means) 19 among others.
The fuel injector 13 is mounted internally of each cylinder of the engine 1 for injecting the fuel directly into the combustion chamber defined within the cylinder. A crank angle sensor 14 for generating a crank angle signal CA is installed in association with a crank shaft which is driven rotatively by the engine 1.
For detecting a depression stroke .alpha. of an accelerator pedal manipulated by an operator or driver of a motor vehicle or the like equipped with the engine system now under consideration, an accelerator pedal stroke sensor 15 is provided in association with the accelerator pedal (not shown).
The crank angle signal CA and the accelerator pedal depression stroke signal .alpha. are inputted to the ECU 12 similarly to the other sensor signals.
As additional sensors, there are provided an intake pressure sensor for detecting the intake air pressure within the intake pipe of the engine 1, an intake-air temperature sensor for detecting the temperature of the intake air and a cooling water temperature sensor for detecting the temperature of the cooling water of the engine among others, although they are not shown in the figure. Parenthetically, the intake pressure sensor also functions as an atmospheric pressure sensor when the engine 1 is stopped.
Additionally, an ISC actuator for controlling the opening degree of an ISC (Idle Speed Control) valve provided in association with a bypass passage of the intake pipe 1a is provided as another actuator (not shown) for the combustion control of the engine 1.
The crank angle sensor 14 outputs a pulse signal corresponding to the engine rotation number (engine speed (rpm)) as the crank angle signal CA so as to serve also as a rotation sensor (engine speed sensor), as well known in the art. Further, the crank angle signal CA contains pulses having edges corresponding to reference crank angles of the plural cylinders, respectively, of the engine wherein each of the reference crank angles is employed for arithmetically determining the control timing for the engine 1.
An exhaust gas recirculation passage (hereinafter referred to as the EGR passage) 16 is provided between the exhaust pipe 1b and the surge tank 5 for recirculating a part of the exhaust gas into the intake pipe 1a,wherein a stepping-motor-driven type EGR regulating valve 17 (EGR regulating means) is provided in association with the EGR passage 16 for regulating the amount or quantity of the exhaust gas recirculated to the intake pipe. This quantity will hereinafter be referred to also as the EGR quantity.
Further, disposed downstream of the surge tank 5 is an intake port which is divided into two sections for each of the engine cylinders, wherein a swirl control valve 18 is provided in one of the intake port sections for controlling the generation of swirls of the air-fuel mixture within the engine cylinder.
More specifically, the swirl control valve 18 is operated by a stepping-motor type actuator 19 to be regulated in respect to the angular position (i.e., opening degree) so that the swirl rate within the engine cylinder can be controlled.
The swirl control valve 18 and the actuator 19 cooperate to constitute the swirl rate regulating means for regulating the swirl rate within the cylinder.
An onboard battery 20 supplies electric power to the ECU 12 by way of an ignition switch 21.
FIG. 24 is a block diagram showing an exemplary configuration of the ECU 12 mentioned previously by reference to FIG. 23. Referring to FIG. 24, the ECU 12 includes a microcomputer 100, a first input interface circuit 101, a second input interface circuit 102, an output interface circuit 104 and a power supply circuit 105.
The first input interface circuit 101 shapes the crank angle signal CA to thereby generate an interrupt signal, which is then inputted to the microcomputer 100.
On the other hand, the second input interface circuit 102 is designed as to fetch the other sensor signals (e.g. signals indicative of the intake air quantity Qa, throttle valve opening degree .theta., an air-fuel ratio F, the accelerator pedal depression stroke .alpha., etc.) as the input signals to the microcomputer 100.
The output interface circuit 104 is designed to amplify the various actuator driving signals (e.g. the intake-air flow control signal A, the ignition signal G, the fuel injection signal J, etc.) to output the amplified signals to the throttle valve actuator 7, the ignitor 11, the fuel injector 13, etc., respectively.
The microcomputer 100 is comprised of a CPU (Central Processing Unit) 200, a counter 201, a timer 202, an A/D (analogue-to-digital) converter 203, a random access memory (hereinafter referred to as the RAM in abbreviation) 205, a read-only memory (hereinafter referred to as the ROM in abbreviation) 206, an output port 207 and a common bus 208.
The CPU 200 serves to arithmetically determine the control quantities for the throttle valve actuator 7 and the fuel injector 13 in dependence on the engine operation state (e.g. the accelerator pedal depression stroke .alpha. and the crank angle signal CA) in accordance with a predetermined program.
The free-running counter 201 is designed to measure a rotation period of the engine 1 on the basis of the crank angle signal CA, while the timer 202 is employed for measuring or determining various control time points or timing and time durations or periods.
The A/D converter 203 converts the analogue input signals from the various sensors to digital signals which are then inputted to the CPU 200.
The RAM 205 is used as a work memory for the CPU 200 while the ROM 206 is used for storing therein various operation programs to be executed by the CPU 200.
Various control signals (e.g. the fuel injection signal J, the ignition signal G, etc.) are outputted through the output port 207. The aforementioned individual components 201, 202, 203, 205, 206 and 207 incorporated in the microcomputer 100 are connected to the CPU 200 by way of the common bus 208.
Next, description will be directed to the operation of the conventional control system for the cylinder injection type internal combustion engine of the structure described above by reference to FIGS. 23 and 24.
In the course of controlling the operation of the engine 1, the signals indicative of the engine operation state (i.e., sensor signals) are inputted to the ECU 12 from the various types of the sensors mentioned previously.
When the crank angle signal CA is inputted to the ECU 12A, an interrupt signal is issued through the first input interface circuit 101 in response to a pulse edge of the crank angle signal CA.
In response to the interrupt signal, the CPU 200 reads the output content or value of the counter 201 to thereby determine arithmetically the rotation period of the engine 1 on the basis of a difference between a current counter value and a preceding one, the rotation period as determined being then stored in the RAM 205. Further, the CPU 200 arithmetically determines the engine rotation number or engine speed Ne (rpm) on the basis of the rotation period and the measured time or period corresponding to a predetermined crank angle which can be derived from the crank angle signal CA.
On the other hand, through the second input interface circuit 102, the analogue sensor signals such as the signal indicative of the accelerator pedal depression stroke .alpha. and others are fetched to be supplied to the CPU 200 after having been converted to the corresponding digital signals by the A/D converter 203.
A control parameter arithmetic means realized by the CPU 200 arithmetically determines various control parameters on the basis of the sensor information indicative of the engine operation states to thereby output driving signals corresponding to the control parameters to the relevant actuators mentioned previously by way of the output port 207 and the output interface circuit 104.
By way of example, the CPU 200 incorporated in the ECU 12 arithmetically determines a desired opening degree of the throttle valve (hereinafter referred to as the desired throttle valve opening degree) on the basis of the accelerator pedal depression stroke .alpha. to output the intake-air flow control signal A corresponding to the desired throttle valve opening degree. In response to this signal A, the throttle valve actuator 7 is so driven that the actual throttle valve opening degree detected by the throttle position sensor 4 coincides with the above-mentioned desired throttle valve opening degree.
Further, the CPU 200 arithmetically determines a desired fuel injection quantity to output the fuel injection signal J corresponding to the desired fuel injection quantity. In response thereto, the fuel injector 13 is actuated for a pulse duration or width which allows the actual fuel injection quantity (determined by the duration of actuation of the fuel injector 13) to coincide with the desired fuel injection quantity and at a predetermined timing based on the crank angle signal CA to inject the fuel directly into the associated cylinder of the engine 1.
Besides, the CPU 200 arithmetically determines a desired ignition timing to output the ignition signal G indicative of the desired ignition timing for thereby driving the ignitor 11 at a predetermined timing in synchronism with a fuel injection timing.
Consequently, the primary current of the ignition coil 10 is interrupted in synchronism with the ignition signal G, whereby the high voltage induced in the secondary winding of the ignition coil 10 is applied to the spark plug 8 through the distributor 9. Thus, electric discharge occurs at the spark plug 8 at a predetermined ignition timing to generate the spark for ignition.
In this manner, by injecting the predetermined amount or quantity of fuel into each cylinder of the engine 1 and by igniting the mixture gas containing the fuel as injected at the predetermined ignition timing, optimal operation of the engine 1 can realized.
In general, the CPU 200 is so designed as to arithmetically determine a desired mean effective pressure which is in proportion to the torque for thereby setting the parameters for combustion or combustion parameters such as the air-fuel ratio, fuel injection timing (fuel injection end timing) and the ignition timing. For more particulars in this respect, reference should be made to Japanese Unexamined Patent Application Publication No. 312433/19996 (JP-A-8-312433).
In other words, the control parameters are arithmetically determined on the basis of the desired mean effective pressure derived as the presumed engine load information. In this conjunction, it is however noted that a time lag will unavoidably be involved in the aforementioned arithmetic determination of the desired mean effective pressure. In that case, the presumed engine load does not coincide with the actual engine load. Consequently, an error is involved in the engine control, which in turn incurs a possibility of degradation of the fuel cost performance and increase of harmful components contained in the engine exhaust gas.
Especially in the transient operation phase of the engine such as accelerating or decelerating operation, the actual quantity Qa of the intake air charged into the cylinder will deviate from the intake air quantity commanded by the accelerator pedal depression stroke .alpha. due to a lag involved in the operation of the throttle valve 3 which is controlled electronically by the throttle valve actuator 7 and/or lag involved in increasing or decreasing the flow rate of the air charged into the intake pipe 1a,whereby the control parameters for the combustion mentioned above deviate from the optimum values, giving rise to a problem that the exhaust gas quality and the fuel cost performance of the engine change for the worse.
Besides, due to the manufactural dispersion or variance in the mechanical structure of the intake system (inclusive of the intake pipe 1a) of the engine 1, the intake air quantity may vary even in the steady or cruising operation thereof. Consequently, the control parameters for combustion mentioned above may deviate from the optimum values, unlike the case where the actual intake air quantity coincides with the one commanded by the acceleration pedal depression stroke signal .alpha..
Furthermore, in the case of the conventional system disclosed in the above-mentioned patent application publication (JP-A-8-312433), the combustion mode of the engine 1 is determined on the basis of the engine rotation number or engine speed (rpm) Ne and the desired mean effective pressure, whereon the control parameters conforming to the determined combustion mode are arithmetically determined.
More specifically, as the combustion modes which are changed over to one another in dependence on the operation state of the engine, there may be mentioned a stratified lean combustion mode for realizing a lean combustion (lean burn) by producing a stratified mixture gas by injecting the fuel during the combustion stroke of the engine, a uniform lean combustion mode for realizing a lean combustion by producing a uniform mixture gas by injecting the fuel during the suction stroke, a stoichiometric combustion mode realized by performing a stoichiometric feedback control, and an open loop mode in which the engine is operated with an air-fuel ratio smaller than the stoichiometric air-fuel ratio.
To this end it is noted that because of the lag involved in the arithmetic determination of the desired mean effective pressure, deviation of the control from the optimal may sometimes occur in the combustion modes mentioned above.
Especially in the lean combustion mode (or lean burn mode), operation of the engine will become unstable when the engine 1 is in the cold state with the temperate of the mcooling water being lower than a predetermined temperature or when the temperature of the intake air is extremely low, as encountered in driving the motor vehicle in a cold district, or when the atmospheric pressure is low, as encountered in driving the vehicle in a highland area (i.e., when the conditions for the combustion are poor), because the lean combustion or burn behavior becomes degraded in these cases.
Furthermore, at the time of changeover of the combustion modes, it is impossible to control the fuel injection timing (fuel injection end timing) and the ignition timing such that change in the torque loss brought about by the pumping loss, cooling loss or by other causes upon changeover of the combustion mode can be coped with. Thus, there may arise a possibility of occurrence of torque shock.
In this conjunction, it is additionally noted that attempt for preventing such torque shock will incur much complexity in the combustion mode changeover control procedure executed by the ECU 12.
Additionally, in the conventional system disclosed in the patent application publication (JP-A-8-312433) mentioned previously, control of the opening degree of the EGR regulating valve is performed on the basis of the aforementioned desired mean effective pressure as well. Consequently, lag due to the arithmetic operation involved in determining the desired mean effective pressure can not be avoided. Besides, the arithmetic operation itself is intrinsically very complex.
More specifically, because of occurrence of the lag in the arithmetic operation for determining the desired mean effective pressure, difficulty is encountered in controlling the EGR quantity conformably to the engine load. Additionally, the EGR quantity may deviate from the optimum value due to manufactural dispersion or tolerance of the throttle valve and change of the environmental conditions.
A method of controlling the swirl control valve 18 and the swirl valve actuator 19 is described, for example, in Japanese Unexamined Patent Application Publication No. 79337/1993 (JP-A-5-79337). According to the teachings disclosed in this publication, the CPU 200 is so programmed as to partition the engine load range into three regions as a function solely of the accelerator pedal depression stroke .alpha., wherein the opening degree of the swirl control valve (i.e., angular position of the swirl control valve 18) is selectively set to a value in one of the three regions in dependence on the engine load.
Nevertheless, the optimum swirl rate conforming to the air-fuel ratio and the engine load can not always be realized, giving rise to the problem that the combustion mode may change for the worse.
As will now be appreciated from the foregoing description, the conventional control system for the cylinder injection type internal combustion engine suffers a problem that error or deviation in the control can not be avoided due to a lag involved in the arithmetic operation for determining the desired mean effective pressure particularly in the transient engine operation phases, incurring degradation in the quality of the exhaust gas as well as the fuel cost performance, because the various control parameters such as the desired air-fuel ratio, the desired fuel injection timing and the desired ignition timing are determined on the basis of the desired mean effective pressure.
Furthermore, even in the cruising or steady operation phase of the engine, the control parameters may depart from the respective optimum values because of the manufactural dispersion in the mechanical structure of the intake system, which leads to a problem that difficulty is encountered in realizing the optimal combustion state.
Additionally, in the conventional control system for the cylinder injection type engine, the combustion mode of the engine is determined on the basis of the desired mean effective pressure and then the control parameters are arithmetically determined so as to conform with the combustion mode as determined. Thus, when the environmental conditions for the engine operation become worse in the lean combustion mode among others, the combustibility or combustion behavior in the lean combustion mode will be degraded because of the lag involved in the arithmetic operation for determining the desired mean effective pressure, which in turn brings about instability of the engine operation, presenting another problem.
Furthermore, in the conventional control system, it is difficult or impossible to control the fuel injection timing (fuel injection end timing) and the ignition timing at the time of changeover of the combustion modes such that change in the torque loss brought about by the pumping loss, cooling loss or by other causes upon changeover of the combustion mode can be compensated for. Thus, there may arise a possibility of occurrence of torque shock. Further, attempt for preventing such torque shock will incur much complexity in the combustion mode changeover control procedure.
Additionally, in the case of the conventional system, control of the opening degree of the EGR regulating valve is also performed on the basis of the desired mean effective pressure. Consequently, the control can not evade a lag brought about by the arithmetic operation for determining the desired mean effective pressure. Besides, the very arithmetic operation for determining the EGR valve opening degree becomes very complex, to another disadvantage.
Moreover, due to occurrence of the lag in the arithmetic operation for determining the desired mean effective pressure, difficulty is encountered in controlling the EGR quantity so as to conform with the engine load. Additionally, the EGR quantity may deviate from the optimum value due to manufactural dispersion or tolerance of the throttle valve and change of the environmental conditions. Thus, there arises a problem that NO.sub.x contained in the exhaust gas can not be removed satisfactorily, incurring degradation of the exhaust gas quality.
Further, in the conventional control system for the cylinder injection type internal combustion engine, the swirl rate (i.e., angular position of the swirl control valve 18) is controlled on the basis of the partitioned engine load regions in dependence on the accelerator pedal depression stroke .alpha.. With such arrangement, the optimum swirl rate can not always be realized, although it depends on the air-fuel ratio and the engine load state, which also leads to the problem that the optimal combustion state can not be realized.